Gas turbine process with addition of steam



May 18, 1954 B. MILLER GAS TURBINE PROCESS WITH ADDITION OF STEAM 3Sheets-Sheei'I l Filed Feb. 21, 1951 ATTORNEY May 18, 1954 B. MILLER GASTURBINE PROCESS WITH ADDITION oF STEAM Filed Feb. 21, 1951 3Sheets-Sheet 2 INVENTOR ATTORNEY B. MILLER May 18, 1954 GAS TURBINEPROCESS WITH ADDITION OF STEAM Filed Feb. 21, 1951 6 Sheets-Sheet 3ATTORNEY Patented May 18, 1954 `GAS TURBINE PROCESS WITH ADDITION orSTEAM Benjamin Miller,-0zone Park, N. Y., assignor to The `ChemicalFoundation, Incorporated, 1a. New York membership corporationlApplication February 21, 1951, Serial No. 212,060

Thisinvention relates to amethod of increas ing `the power :output ofcontinuous ow Vlgas turbines.

' The `principal :purpose ofthe invention is to provide novel methodsand apparatus for `producing power .by the expansion through turbines ofcompressed fgases :heated by internal combustion, such novel methods andapparatus being especially :adapted to `reduce the `cost per `horsepower:of `the machines.. as` compared' to those of `current types,whenfdesigned .to obtain comparafble thermal efiicienoy. l l .Thetypicalinternal combustion turbine cornprises essentiallythree elements."The iirst element is a compressor, which takes in atmospheric Aairandforces it under Asubstantial supera-'omospheric fpressure "into thesecond element, a com `bustion :chamber `or .combustor in which the `airmeets fuel :and combustion takes Lplace, Aproducing hot gaseswhich'passto andthroughlthe third element, a turbine, in whichLexpansionof` the hot gases takes place with the generation ofmechanical energy, part of which drives the lcompressor while 'thebalance is available to drive a load.

The invention can be best` appreciated and A:evaluated by considering`first the operation of conventional apparatus now employed. As willFb'e :explained hereafter, the .novel apparatus of the invention, forequal thermal eiciency pro- -duces approximately twice` as much". poweras currently employed `.units of about .thefsame `size :and operatingunder the same `temperature and pressure conditions; this reduces thecost per horsepower 'to 40% less than that yof` comparable `:prior :artunits.

In `a typical `internal combustion turbine currently employed, `an Aair'flow rate of 90 pounds perv second is used to produce 6,140 horsepowerfor the vload with .a thermal eiciencyo'f 18.2 per cent. `The air entersthe intake duct `at `80 E'. and 14.1 pounds per square inch absolute ip.rs. i. a), assuming an altitude `of. 1000 ft., and suffers a pressureAdrop oi 0.2 p. s. i. a. in passing through the `air washer and intakeduct, reaching thecompressor at 80 'F. `and 13.9 p. rs. i. a. Theaxial-now compressor having an efliciency of 8B `per cent raises thepressure to "76."5 p. s. i. la. and the temperature to 470 F., absorbing12,050 horsepower and putting 30.73

.million `B. t, `u. per hour into the. 11,192 moles i? 10 Claims. (Cl.(S0-39.05)

2 moles per hour, which is'equivalent to 35.95 million B. t.u. (lowerheating value) per hour. The combustion is not perfect, and there issomeheat lost to the surroundings, so that the enthalpy increase is only83.45 million B. t. u. per hour, and the hot gases reach `the turbine at1400 F. and r71.8 p. s. i. a., the pressure loss throughthe combustorand connectionsbeing 4.7 p.\s.i. a.`

In passing through the turbine, lthe pressure falls `to 14.2 p, s. i.a., and the temperature to 868 the eciency of the turbine being 85 percent. The enthalpy drop is 48.11 million B. tfu. per hour, of which 1.73million B. t. u. per hour are lost to the surroundings,1includii'igdrect loss and frictionin the turbine andcompressor. The

balance, 46.38 million B. t. u. per hour, is-convertedinto usedmechanical energy; this is 18,'190 horsepower, of which the compressortakes 12,050

fto return to `the air, Vwhile .6,140 horsepower remain for the load.

The thermal efficiency `is 18.2 per cent; the combustor loss is 2.9 percent, and the .losses `in turbine and compressor total 2.0 per cent,while '76;9 per cent of the heatof the fuel remains `as sensible heat inthe gases `which leave Vthe turbine at 868 F. and 14.1 p. s. i. a., thepressure drop through the exhaust duct being0.1 p. s. i. a.

Some of the heatin the exhaust'gases may be used. in various ways, asfor heating water or generating `low-pressure steam; a part maybe`returned to the cycle by means of an indirect `heat exchanger, orregenerator, having two flow paths separated by `a diaphragm, one pathbeing inserted between the turbine `exhaust, and the stack, 'and theother path betweenthe compresso discharge and the combustor.

The fraction of the `heat in the exhaust gases which can be returned `tothe cycle by means of a regenerator depends upon the temperature atwhich the gases `leave the turbine, the temperature at which the airleaves the compressor, and the number of square feet of heat transfersurface in the regenerator. With a regenerator having `50,000 squarefeet of heat transfer surface addedto the equipment described abovesuicient heat can be transferred from the exhaust gases to the air to raiseVthe thermal efficiency to 25.0 per cent, but only at the expense of asubstantial decrease in power output, from 6140 horsepower to 5550horsepower.

The vdecrease in power output is caused chiefly by the added pressuredrops which comewith `the insertion of the regenerator. The addedpressure drop between the 'compressor and the combustor acts to decreasethe pressure at `the turbine inlet from 71.8 p. s. i. a. to 70.5 p. s.i. a.. While the added pressure drop between the turbine exhaust and thestack acts to increase the pressure at the turbine exhaust from 14.2 p.s. i. a. to 14.6 p. s. i. a. Thus the entropy rise due to the pressuredrop, which is proportional to the logarithm of the ratio of inletpressure to exhaust pressure, is made less, and the correspondingenthalpy decrease is made less also. In this case the temperature at theturbine exhaust becomes 880 F. rather than 868 F., which decreases theturbines used power output from 18,190 horsepower to 17,600; since theair flow rate and the pressure at the compressor discharge is notchanged, the power taken'by the compressor is not changed either, sothat all of the decrease in turbine output is suiered by the'load.

The fuel rate is decreased from 85.95 million B. t. u. per hour to 56.61million B. t. u. per hour,

or from 249 to 164 moles of methane per hour.

The decrease in total moles through the turbine from 11,441 Vto 11,356is partially responsible for the decrease in power output, but theprincipal cause is the decrease in the ratio of inlet pressure toexhaust pressure.

The regenerator adds about per cent to the cost of the gas turbine andreduces its power output Aabout 10 per cent, so that the increase inthermal efliciency from 18.2 per cent to 25.0 percent is accomplished atthe expense of an increase of about one-third in cost per horsepower ofoutput.

By increasing the temperature of the gases at the turbine inlet bothpower output and thermal efliciency can be increased. However, at thepresent time the practical maximum temperature is 1500 F., this limitbeing imposed by materials of construction, particularly for the turbinerotor.

With the turbine inlet temperature limited to 1500 F., increased thermalefciency can be obtained by increasing the complexity and cost of theapparatus.

InY one such suggested method the compression is carriedY out in twostages, with an intercooler between the two stages. pression is therebyreduced for the same pressure rise, or a greater pressure rise may beobtained with the same work. The heat abstracted by the intercoolercannot in general be recovered, but a large part of it can be replacedbythe regenerator, since the temperature of the air leaving 'thecompressor is less than it would otherwise have been, and the lower thetemperature of the 4air leaving the compressor, the greater the quantityof heat which the regenerator can transfer to it.

In a practical operation of such a plural compressor unit, an air flowrate of 72 pounds per second is used tc produce 7,130 horsepower for theload with a thermal eiiiciency of 29.8 per cent. The air enters theintake duct at 80 F. and 13.9 p. s. i. a., the altitude being 1500lfeet, and suffers a pressure drop of 0.2 p. s. i. a. in passing throughthe air washer and intake duct, reaching the first compressor at 80 F.and 13.7 p.V s. i. a. The 86 per cent eiiioient rst compressor raisesthe pressure to 42.3 p. s. i. a. and the temperature to 316 F.,absorbing 5800 horsepower and putting 14.78 million B. t. u. per hourinto the 8,953 moles of air.

The air then iiows through an indirect heat exchanger, or intercooler,where it gives up 13.84 million B. t. u. per hour in cooling. to 95 F.The heat is absorbed by water ilowing through the intercooler at therate of 800 gallons per minute The work of com- 4 and being heated fromF. to 105 F. The intercooler has 21,400 square feet of heat exchangesurface.

In passing through the intercooler and ducts the air suiers a pressuredrop of 1.4 p. s. i. a., and enters the second compressor at 40.9 p. s.i. a. and 95 F. The 86 per cent efficient second compressor raises thepressure to 131.3 p. s. i. a. and the temperature to 347 F., absorbing6180 horsepower and putting 15.77 million B. t. u. per hour into theair. Y

From the second compressor the air flows through the regenerator to thecombustor. The regenerator, which has 26,200 square feet of heatexchange surface, puts 22.23 million B. t. u. per hour into the airraising its temperature to 690 F. In the combustor the air meets fuel,supplied at the rate of 61.09. million B. t. u. (lower heating value)per hour, or 177 moles of methane per hour. There is a loss of 1.60million B. t. u. per hour in the combustor, and a pressure drop'of 11.2p. s. i. a. through the regenerator, the ducts, and the combustor, sothe gases at the rate of 9,130 moles per hour reach the turbine at 1500F. and 120.1 p. s. i. a.

In passing through the turbine the pressure falls to 14.4 p. s. i. a.and the temperature to 850 F., the efficiency of the turbine being percent. The enthalpy drop is 50.24 million B. t. u. per hour, of which1.51 million B. t. u. per hour are lost, and 48.73 million B. t. u. perhour are converted to used mechanical energy; this is 19,110 horsepower,of which the compressors take 11,980 to return to the air, leaving 7130horsepower for the load.

From the turbine the gases pass through the regenerator and the exhaustduct to the stack, which they reach with a temperature of 475 F. and apressure of 13.9 p. s. i. a., having given up 22.23 million B. t.' u.per hour to the air passing to the combustor, and still carrying 25.83million B. t; u. per hour. I

Thus the total heat of 61.09 million'B. t. u. per hour supplied as fuelis accounted for as follows: 18.18 million or 29.8% to the load; 25.83million to the'stack; 13.84 to the cooling water; 1.60 lost by thecombustor; 1.51 lost by the turbineand compressor, and 0.13 lost by theregenerator and ducts.

By adding a second intercooler, a thirdv compressor, a second combustorand a second turbine-the power output per pound of air handled and thethermal eiliciency can be further increased. Such plural compressorplural combustor cycle produces 13,240 horsepower with a How rate of 72pounds of air per second at a thermal eiiciency of 35.3 per cent. Thesecond intercooler, like the first, has 21,400 square 'feet of surface,and it cools the air from 347 F. to 98 F., removing 15.59 million B. t.u. per hour from the air and heating 780 gallons of water per minutefrom 70 F. to 110 F. In passing through the second intercooler and theducts the pressure drops 4.3 p. s. i. a., so the air enters the thirdcompressor at 127 p. s. i. a. and 98 F. The 86 per cent efficientcompressor raises the pressure to 331 p. s. i. a. and the temperature to336 absorbing 5850 horsepower and putting 14.92 million B. t. u. perhour into the air.

From the third compressor the air iiows through the regenerator to thefirst combustor. The regenerator, which has 39,300 square feet of heattransfer surface, puts 37.58 million B. t. u. per hour into the air,raising its temperature to 906 F. Burning of 132 moles of methane per'spondingly The steam may be generated at a higher pressure andtemperature, or it may be superheated in an unnred superheater, or theexhaust gases may pass, through a regenerator which heats the air fromthe compressor and then through the unred boiler. For example,generating the steam at 320 F. instead of 310 F. would increase thethermal efficiency to 25.0 per cent, but would require more heatexchange surface, particularly in the water heater.

Comparing the cycle and apparatus of Fig. 1 with prior installationsemploying a single compressor with a compressed air preheater, it may beseen that the novel cycle of Fig. 1 produces 117 per cent more power atabout the same thermal efficiency with the same air compressor, butpasses about 34 per cent more volume through the combustor and turbine,requires the turbine to generate and transmit about 37 per cent morepower, and requires about 25 per cent less heat transfer surface. Theadditional cost considering combustor, turbine, and heat exchangesurface is about 20 per cent, so that the cost per horsepower isdecreased by more than 40 per cent.

It will be noted that the boiler 9 and water lheater l of Fig. 1 haveeach but one path in the gas circuit. Thus the boiler and water heatercan be designed for low pressure drop in the gas circuit and highpressure drop in the water-steam circuit; this improves the power outputof the turbine with but a small additional load on the feed water pump,which in any event takes only a few horsepower. It will be noted alsothat as compared to prior installations where the maximum temperature ofthe heat transfer surface in the regenerator is comparable to thetempera ture at which the gases leave the turbine, in the novel systemof the invention the maximum temperature of the heat transfer surface inthe boiler 9 is' not much greater than the temperature of the boilingfluid.' By reason of the high allowable pressure drop in the water-steamilow path and the lower maximum temperature, the cost of the boiler andwater heater, per square foot of surface, can be substantially less thanthat of the regenerators used in prior art systems.

The division of the compressor into two or more stages, withintercooling between the stages and the division of the turbine intostages of expansion, with reheat between stages, can be employed toimprove thermal eniciency. It has been found desirable, particularly inconnection with the improved cycle, to use adiabatic cooling betweencompressor stages, the cooling being accomplished by the injection ofliquid water which vaporizes.

The first of these improvements is illustratedI lby the flow diagram ofFig. 2. As there shown the system comprises the rst compressors 2l andsecond compressor 22 mounted on the shaft 23. The turbine 24 similarlyis mounted on shaft 22 and drives the two compressors and load 25. As inthe rst described embodiment, air is taken from an air supply 26 and ispassed through the air washer 21 and thence to the rlrst compressor 2l.Air compressed in the first compressor as 4shown is passed through theadiabatic cooler and thence through the second compressor 22 to thecombustor 29. The cooler 28 preferably is a spray chamber type; this ismuch less costly than an indirect cooler and imposes a much smallerpressure drop than an indirect cooler although it may not be quite soeffective in reducing the Awork of the compressor 22.

AS in the embodiment illustrated in Fig. 1,

steam is generated in a steam generator, such as the unred boiler 30,byv vaporizingpreheated water by indirect heat exchange with hot exhaustgases from the turbine. These gases thus pass to the water heater 3l topreheat the water supplied to steam generator and are discharged to thestack.

In the combustor the air from the second compressor in admixture withsteam from generator 30 meets the stream of fuel from fuel supply 32 andthe products of combustion are expanded in the turbine 24.

In such a cycle the air flow rate to the -compressor is 72 lbs. persecond. The cooler 28 is designed to introduce 466 moles of water perhour, and the mixture of air and water vapor enters the secondcompressor at 170 F. and 41.7 p. s. i. a. The second compressor raisesthe pressure to 133.8 p. s. i. a. and the temperature to 456 F.absorbing 7530 horsepower and introducing 19.31 million Bf t. u. perhour into the air and steam.

In the combustor the air-steam mixture from compressor 22 meets fuelflowing at the rate of 109.2 million B. t. u. (317 moles of methane) perhour, and steam from generator 30 flowing at the rate ofi-'2334 molesper hour. There is a loss of 2.72 million B. t. u. per hour, and themixture of combustion products and added steam enter the turbine 24 at1500 F. The pressure drop in the combustor and connections is 8.4 p. s.i. a., so that the pressure at the turbine inlet is 125.4 p. s. i. a.

The total flow through the turbine is 12,070 moles per hour, and theturbine exhaust pressure is 14.4 p. s. i. a.; the turbine einciencybeing 85 per cent, the turbine exhaust temperature is 822 F.

The enthalpy drop is 68.35 million B. t. u. per hour, of which 1.88million B. t. u. per hour are lost, the remaining 66.47 million B. t. u.per hour being converted into used mechanical energy; this is 26,070horsepower, of which the two comu pressors take 13,350 horsepower toreturn to the air, while 12,740 horsepower remain for the load. Thus thethermal efficiency is 29.7 percent.

From the turbine the gases pass, asexplained, through the unred boiler39 having 17,200 square feet of heat exchange surface, the unred waterheater 3l having 12,000 square feet of heat ex` change surface, and theexhaust duct to the stack, which they reach at a pressure of 13.9 p. s.i. a. and a temperature of 293 F.

The boiler 30 generates at 3507F. and 134.6 p. s. i. a. the 2334 molesof steam per hour which are added to the circuit at the combustor, using38.07 million B. t. u. per hour and cooling'the gases from 822 F. to 420F. The water heater provides the boiler feed, heating 2334 moles ofwater per hour from F. to 350 F., and using 11.43 million B. t. u. perhour.

Comparing the cycle of Fig. 2 with the plural compressor and intercoolersystems of the prior art, the improved cycle produces about 78 per centmore power at aboutV the same air rate, using a second stage compressorwhich requires about 22 per cent more power because it handles about 5per cent more moles taken in at about 14 per cent higher absolutetemperature, a combustor and a turbine which handle about 32 per centmore moles, a turbine which generates and transmits about 37 per centmore power and an additional spray chamber, but using about 38 per centless heat exchange surface. The additional cost of making the secondcompressor, combustor amasar and: turbine slightly larger about balancesthe saving on` heat exchange equipment, so that the cost per `horsepowershould* be decreased more than 40 per cent.

. A further imodication of the invention which is` shown.diagrammatically in Fig. 3 involves the use. of.` multiple stages ofcompressors and tur bines withl reheatingl of.` the gas` between the twoturbine: or expansion stages. 3 `the compression is. divided lntoi`three stages comprising. the compressors 441., 42 and 43, eachy mountedon` the` shaft 44 which` shaftn isvdriven by the turbines 45v and 46?.`Power. is transmitted from` the` turbine to shaft; 641` to. drive thecompressors and also the load 4T.. Asoompared with` prior artinstallations using an equivalent com-y bination of pluralcompressorsand turbines, but without utilizing thenovel cooling conceptof the invention, theinstallation depicted in. Fig. 3 producesl109 percent` more power with about the same thermall eiiciency.

Thiamodiiication, as indicated in the` drawings. utilizes `intercoolingbetween the sequential4 compression `stages and reheating' between the`expansion stages. The operation willl have been. appreciated from theearlier discussion of` the first two` modifications. Air from air`supply 418 is washed in washer 49 and is passed to the first compressorML. Compressed airffrom compressor 4l isthen cooled. inthetadiabaticcooler 50and is passed sequentially through the second':com-

pressorv42 and cooler 51.` to the third compressor 43. The` mixture ofcompressed air and `Water. vapor from` compressor 43` as` shown iscontinu*- ously chargedto the iirsthcombustor 52.

As` in.` the previously described drawing, steam isfad'mitte'di to thecombustion zone to `effect the desired cooling. of the combustion gases.Such steam is.A formed in a. suitable steam generator such as theunredboiler 53T, the heat for this generator being supplied as shown',by indirect heat exchange withhot gases from the second turbine 4.6'.These exhaust gases are further cooled in the water heater 514` whilepreheating the water fedv to boiler53'..

In.'` the rst combustor 52 the mixtureV `of air and steam meetstheifuell fed fronriuel1 supply 55 `and the products. of', combustion`as shown are expanded inY turbine f 45. From'. the: rst' turbinetheexpanded gases pass.` to the. second combustion 561fed1s with `fuelfrorrrsupply 55; and` .the products of combustion are'- expanded inthe.v secondi turbine 46. Theturbine exhaust-as notedrisxutilizedtogen'erate. steam'` and. preh'eat` the water `supply. to. the steam4generator thus most effectively loweringrthe stack temperature. of the`exhaust gases.

This installation employs 72 pounds of. air: per second to `produce27,710 `horsepowerV with 3.5.6 per .cent1 thermal'. eiciency to produce,as; previa i ouslynoted,. 109 per cent more powerwithabout theV samethermal. eiiiciency as; presently utilized installations of theequivalent 'plural` compression and; turbine stages.

Gompressors 4l and 42 and thei irst` intercool'er .5 0L are'` identical.with.' those previously dezscribed` 2,; Ysimilarly the. inietlairconditions andthe: conditions at the; `discharge of;secondicompressor.AZaraidentical withtthoseo Eig. 2.V The. secondadiabatic'. cooler 5I introduces -650 moles of water` vapor per hourinto the circuit concomitantly reducing-the.l temperature toi 260 F'.This second intercooler andconnections con.- sumefliapLs. i; aandthezngasesenterfthethird compressor. at` .132 gp. s.. i.V a: The.thirdieompres- As4 shown in 10 sor raises the. pressure to 396 p. s.. i.a. and the temperature to 553 F. absorbing 6430 horse.- power andputting. 21.51 million B. t. u. into the mixture. off air and steam.

The mixture of steam and air passesfrom the compressor' 43 to the.combustor 52 andis admixed with fuel introduced at the rate of 124.25million B. t.` u. (350 moles of methane) per hour and steam entering`from generator 53 at the. rate or" 4000 moles per hour. The thermal lossin combustor 52. is 3.18 million B. t. u. per hour,xand the mixture` ofcombustion products and steam enters turbine i5 at l500 F. During`passage through the combustor 52 and its connections. the pressure drop4is 25 p. s. i. a..,andihence the pressure-attire inletof turbine 45. is`371 p; s; i. a. The gases: expand in turbine 45 down. to 69.5 p. s. i.a.; the turbine efliciency being 'per cent, thetemperature ofithe gasesexhausted from tur.- bine45 is 962 F.

in the. second combustor 55 fuel is added at the rate of 73.75 million.B. t. u. (214. moles oi' methane) per hour.` The loss in the secondcombustor is` 2.16 million B. t. u. per hour and the gases enterthesecond turbine at 1500 F; The pressure` drop` through the secondcombustor and connections is `5.4. p. s. i. a. and thus: the pressure ofthe gases at theinlet to the second turbineis 64.1 p. s. i. a.

The gases. expand turbine 46 down to 14A p. s. i. a.; with the turbineefficiency 85 per cent, the: temperature of the exhaust gases fromturbine. 46. is 1021 F.

The enthalpy drop inthe turbine; 45 is 67.45 millionB.` t. u. per hour,of which 1.85 million B. t. u. per hour are lost; the enthalpy drop inturbine 45: is: 62.32 million B. t. u. per hour.. of which 1.76 millionB. t. u. per hour are lost; thus the two turbines convert intousedmechanical energy 126.16 million B. t. u; perhoulr, which isie9g470horseposer. The three compressors recycle 21.,760 horsepower to theair,: leaving 27,710 horsepower. for the'load, so that the thermal4eiii.- ciency is 35.6 per cent.

From turbine `13.6, asf noted', the gases iiow throughftheunred boiler53;, which has` 26,700 square feet `of heat.. exchange surface,andrgencrates at. p. s.. i. a. and 445 F. the.v 72,000 pounds of' steamperl hour added at therst combustor, and transfers 60.61 million B..t.."u. per hour` from the gases to the-boiling liquid` While cooling the.gases to 508 F.

From' the unred boiler the gases iiow` through the' unred water heater541 and exhaust duct to theI stack. The vwater heater` has 32,500 squarefeet` of heat exchange surface, and it heats from .80 F. to 445 F. the72,000 poundgof water per hour which` the boiler converts. to steam,putting 27.07l millionB. t. u. per hourinto thewater andcoolingthexgases to 260- F.

Inpassing through the boiler; Water heater and exhaust duct the gases`suier a'. pressure drop of` 0.5 p. s. i. a., `and:y reach. the stack'at13.9` p. s. i. a.

Toproduce 1'09: per cent more. powerthan the comparable prior artmultiple: compressor-mul# tiple turbinef units with the same air'owfrate andthe same thermal eiliciency, the. cycle de.-` picted inFig. 3`requires 22 per.` cent more coin'- pressor` horsepower, combustors. andturbines which handle 159 per cent more moles-of gas,.and a.. turbinewhich `generates and` transmits@ .'59 per centmore powenrand. two"adiabatic intercoolers; but it requires 38 per cent lessaheatexchangesurface, yand none of` thefheat. exchange surface 11 has to operate athigh temperature in the cycle of Fig. 3, while part of the heat exchangesurface in the comparable prior art cycle must operate at temperatureswhich are high for ordinary steel.

It will be appreciated that various mechanical modifications may be madewithout departing from the spirit of the invention. For ex-k ample, eachturbine may be divided into two parts, each with its own shaft, inseries, and these may operate at different speeds; or the stream of gaspassing to a combustor may be split into two or more streams, each ofwhich may then pass to a separate combustor and turbine. Also, while, asdescribed, it is preferred to mix the added steam with the combustionproducts before the turbine inlet, it may be arranged that combustionproducts are supplied to some of the turbinenozzles and steam to theothers, so that the mixing takes place in the turbine.

The usual methods of controlling the speed and load may be used with theapparatus of this invention, but in addition the quantity of steam addedmay be decreased at part load; when the quantity of steam added isdecreased, the pressureand temperature at which thesteam is added. maybe increased, thereby improving the f thermal eiciency at part load.

The water used in the adiabatic intercoolers is preferably free fromnon-volatile matter.; however, in some cases water containingnon-volatiles in solution may be used, provided an excess of water isintroduced, and the unevaporated excess in allowed to flow to wastecarrying the non-volatiles away with it. For marine application theboiler may be fed with sea water, in which case there must be suilcientexcess feed to carry away the non-volatiles.

The heatV remaining in the gases leaving the water heater may be usedfor heating water, or for comfort heating, or `for. preparing distilledwater in evaporators, which distilled water may then be used for theadiabatic coolers or the boiler.

The applications of the invention described above have been for poweroutputs of thousands of horsepower. The invention may also be used forlower power outputs, with proportionately greater benefits. Theaxial-flow compressor is not so. suitable for low iiow rates, such aswould be needed for a power output of less thanone thousand horsepower,as is the centrifugal com.- pressor, but the efficiency of thecentrifugal compressor is substantially lower. Since in the improvedprocess of the present invention the power .taken by the compressor isonly about half of that generated by the turbine when the efficiency ofthe compressor is 86 per cent, whereas in the conventional gas turbinethe 86 per cent efficient compressor takes about two thirds of thegenerated power, the present invention makes it practical to operatewith a compressor having only-60 Vper cent efficiency. Thus the internalcombustion turbine is made practical for poweroutputs of the orderof 200horsepower, with a thermal efciency of 15 per cent.

The improvements obtained by the present invention require for fullexploitation the use of about 10 pounds of water per poundof fuel,`which could be a limiting feature in some cases where water is scarceor for mobile applications where the water must be carried. In suchcases partial advantage may be. obtained by using a smaller quantity ofwater.

It is particularly to be observed that the fundamental concept of theinvention is the use ofsteam as a coolant and the generation vof atleast part of theV steam with heat which could not otherwise be madeavailable for conversion into mechanical energy anyway. In the earlierpe-l riod of the development of the gas turbine many suggestions wereadvanced as to driving the turbine by a mixture of gas and steam. Thesesug` gestions were advanced at a time when the gas turbine then knownhad difficulty driving its own compressor. In the modern successful gastur-A bine the criteria are maximum simplicity in design and maximumreliability in operation.

While the conversion of water intosteam under pressure has the advantagethat it requires nov mechanical energy, exceptfor the little demandedfor the feed water pump, this must be paid for with latent heat and asteam power plant cannot work at high thermal efficiency except byoperating at high pressure. On the other hand, the modern successful gasturbine operates at moderate pressure and high temperature. Hence, inthe past, substituting steam for air as a cooling medium could only bejustified when the thermal ef-Y ficiency of the gas turbine which usedthe lsteam was less than about 12 per cent.

Using the heat in the exhaust gas from the turbine to generate the steamimproves the efciency, to be sure, but the heat in the exhaust gas canalso be used to preheat the air. It makes no difference from the fuelconsumption standpoint whether fuel is burned to generate steam in thecombustion chamber to mix with air preheated by the exhaust gases, orfuel is burned in the combustion chamber to heat cold air which is thenmixed with steam generated by heat abstracted from the exhaustgases-provided the. amount of heat recycled from the exhaust gases isthe same in both cases. But the amount of heat recycled from the exhaustgases need not be the same in the two cases.

If the cost of the equipment were unimportant, the compression of theair could take place at constant temperature, or at least withsubstantially no increase in temperature, and then the amount of heattransferred from the exhaust gases to the Vcompressed air could besuiiicient to increase the temperature of the compressed air almost butnot quite up to the temperature at which the exhaust gases leave theturbine; in such a case the temperature of the exhaust gases would bereduced almost to atmospheric.

However, the cost of the equipment is important. It is therefore noteconomic to attempt to obtain very low final compressor dischargetemperature, nor to increase the temperature. of the compressed air bytransfer from the exhaust gases beyond a temperature still substantiallylower than that at which the exhaust leaves lthe turbine. In theexamples given of gasturbne processes which are representative of thepresent state of the art, the lowest temperature at which the exhaustgases leave the regenerator is 475 F. In this case the amount of heatcarried to the stack by the exhaust gases is over 40 per cent of thatsupplied as fuel. This is heatwhich theoretically could be recovered,but practically can not be, in the currently employed process.

By the present invention, however, a substantial fraction of this heatcan be recovered and put to use in the power generation cycle. Thisrecovered heat may be used to improve the thermal eiciency, orto make upfor theloss which takes place when steam is generated by heat whichcould have. been used for heating the amasar:

13 air. In the il'rst described embodimentI thelobjective is lowermachine cost with the same ther-mal ei'ci'ency, though. the possibilityof improved thermal efficiencyis` mentioned. In the second embodimentthe objective is higher thermal eiciency, without excessive machinecost.

Wliilethe reduction in cost -perzhorsepower will'l be less if thequantity of'water used is less, there willbepossibleian offsetting gainin thermal eiliciency. rlhi-swgainiin thermal:` efficiency may beobtained by Vone or more ofthe expediente previously mentioned; vthatis, by generating the steam at a `higher' temperature and pressure; orby superheating it in an unflred superheater through which the exhaustgases `now to the boiler: orby passing' the exhaust gases through aregenerator; which heats vthe air frornthe compressor, and then throughthe boiler.

`Itwill be understood thatthe ratio of water to air which is to be `usedin eachl case is determined by the form in which the `benefits of theinvention areto beI obtained, provided suficient water is available.With arelatively `small ratio of' water to `air both thermal efficiencyand machine cost are improved. As the ratio of water to air isincreased, the cost per horsepower decreases more andmore, but thethermal efficiency increases to amaximum, and then decreases. In theexamples given the water-air ratio has been chosen `to obtain low costper horsepower while keeping the thermal efficiency aboutthe same as inthe comparable cycles of .the present. stateof `the art. Still lowercost per horsepower may be obtained by employing higher .watervairratios, but thethermal efficiency will be less.`

Inapplying the invention to the improvement ofz a particular gasturbinecycle, the temperature at...whch` the exhaust gases pass to thestack is an. important guide. Thus in the example, given in connectionwith one prior art installation the exhaust gases passto the stack at atemperature of 868*F..carrying 76.9 per cent of the heat of the-fuel,and lthe thermal `efficiency is 18.2 `per cent. In the example given inconnection with prior methods using preheated air the thermal efficiencyis 25.0 per cent, the improvement being due to the regenerator whichreduces the temperature of the exhaust gases to 552 F. and returns tothe process enough heat to raise the temperature of the air from 470 F.`to 800 F. In the example of the invention given in connection with Fig.1 the thermal emciency is 24.8 per cent, the improvement over thecomparable prior art installation being due to the boiler and waterheater which reduce the teinperature of the exhaust gases to 254 F. Inthese examples, therefore, the use of the present invention to obtain alarge decrease in `cost per horsepower with about the same thermaleniciency employs such a Water-air ratio as to decrease the exhaust gastemperature to 254 F. in comparison to the decrease in exhaust gastemperature to 552 F. which takes place when a regenerator is used.

It will be noted that in the representative prior method using multiplecompressors with regenerative heating of air the exhaust gases pass tothe stack at 475 F. while in the comparable cycle of the invention shownin Fig. 2 the exhaust gases pass tothe stack at `293 F.; also, that inthe prior art using multiple compressors and multiple turbines theexhaust gases pass to the stack at 480 F. while in the comparable. cycleof Fig. 3` the exhaust gasespasdl to the stack at` 260 F.v Thus `in `allexamples the present invention is applied to bringA the; temperature ofthe exhaust gases. to a value sub-' stantially lower thanregeneratorscould, butstill substantially higher than atmospheric; l

While a larger regenerator would allow more heat to` be recovered, it.is not possiblefin any casefor a regenerator to reduce the temperatureof thetexhaust gases below the temperature at which the air leaves thecompressor, this: temperature being as high as 470 F. in these examples,while by using a large water-air` ratio the temperature of the exhaustgases. can be brought close to atmospheric, thereby decreasing thesensible heat loss. While the power outputincreases with the water-airratio, the latent` heat loss increases also. It is therefore anfimportant feature of the invention to employ Tsuch a water-air ratioand so much heat transfer surface in each case as to obtain the: desiredbalance between thermal efficiency and cost of equipment, controllingthe thermal efliciencyby the balance between latent heat loss andsensible heat loss and the equipment cost by the balance between heattransfer surface cost and rotating equipment cost. The heat transfersurface may be distributed among water heater, boiler, superheater, `andregenerator, depending upon the water-air ratio employed and therelativecost. of the various forms.

While preferred embodiments of the invention have been described, it isto be understood that these are given to illustrate the underlyingprinciples of the invention and not as limiting its useful scope exceptas such limitations are imposed by the appended claims.

I claim:

l. A power production process comprising the steps of generating steamfrom liquid water; compressing atmospheric air to a temperaturesubstantially superatmospheric; passing the compressed air to a zonewhere fuel is introduced and combustion takes place continuously, butwithout increase of pressure; passing the combustion products and excessair in admixture with the steam continuously through an expansion zonewhere mechanical energy is abstracted, in amount substantially greaterthan the mechanical energy absorbed in the compression step; and passingthe expanded mixture in indirect heat exchange relationship with thewater to generate the steam, the pressure at which the steam isgenerated being not less than the pressure at the entrance to theexpansion zone, and the quantity of heat abstracted from the expandedmixture for recycling to the expansion zone being large enough to reducethe temperature of the expanded mixture to a value :not substantiallygreater than the temperature of the air at the end of the compressionstep whereby the heat absorbed as latent heat in the steam generationstep is supplied by abstraction of sensible heat from said expandedmixture without requiring the combustion of additional fuel to supplysuch latent heat.

2. The process of claim l wherein the steam is superheated by heatabstracted from the expanded mixture, and the expanded mixture passes inindirect heat exchange relation with the steam through the steamsuperheating zone before said mixture enters the steam generation zone.

3. The process of claim 1 wherein the water is heated to the saturationtemprature from a temperature substantiallylower by means of heatabstracted from the expanded mixture whereby the temperature of theexpanded mixture is reduced to a value substantially lower than thetemperature of the air at the end of the compression step.

4. The process of claim 1 wherein the pressure at which the steam isgenerated is made substantially higher than the pressure at theentranceto the expansion zone, whereby the waterair ratio is controlledto increase theV thermal efficiency.

5. The process of claim 1, wherein are comprised the steps ofcompressing the air in a plurality of stages and expansion takes placein a series of zones through which combustion products pass sequentiallyof transferring at least a substantial portion of the heat removed fromthe compressed air during the cooling between stages and at least asubstantial portion of the sensible heat of the gases after theabstraction of mechanical energy therefrom to water to generate steamtherefrom at substantially superatmospheric pressure, and adding saidsteam to the mixture of combustion products and excess air before theexpansion. Y

'6. The process of claim 5 wherein there are employed a plurality ofcombustion zones, arranged sequentially, in each of which fuel isintroduced, With an expansion zone between two consecutive combustionzones, and an expansion zone after the last combustion zone in the flowpath.

7. The process of claim 5 wherein the expansion zone in which the steamis admixed with combustion products is not the first in the series.

8. The process of claim 1 wherein the quantity of steam admixed with thecombustion products and excess air is decreased as the power required isdecreased.

9. The process of claim 8 wherein the temperature of the steam beforeadmixture is increased as the power required is decreased.

16 10. A power production process comprising the steps of continuouslygenerating saturated steam from liquid water compressing air to atempera- `ture substantially superatmospheric; continuously passing thecompressed air to a zone where fuel is introduced and combustion takesplace continuously, but Without increase of pressure, the rate of fuelintroduction being so selected with respect to the rate at whichcompressed air enters said combustion Zone that the ratio of air to fuelin the feed tothe combustion zone is substantially greater thanstoichiometric; continuously passing the combustion products and excessair in admixture with the steam through an expansion zone wheremechanical energy is abstracted, in amount substantially greater thanthe mechanical energy absorbed in the compression step; and passing theexpanded mixture in indirect heat exchange relationship with the waterto generate the steam, the pressure at which the steam is generatedbeing not less than the pressure at the entrance to the expansion zone,whereby the heat absorbed as latent heat in the steam generation step issupplied by abstraction of sensible heat from said expanded mixturewithout requiring thecombustion of additional fuel to supply such latentheat. Y

References Cited in the le of this patent UNITED STATES PATENTS NumberName Date 884,821 Zoelly Sept. 3, 1907 996,324 deFerranti June 27, 19111,400,813 Graemiger Dec. 20, 1921 1,887,001 v Zetterberg Nov. 8, 19321,982,664 Holzwarth Dec. 4, 1934 2,078,958 Lysholm May 4, 1937 2,244,467Lysholrn June 3, 1941 2,469,678 Wyman May 10, 1949 2,549,819 Kane Apr.24, `1951 FOREIGN PATENTS Number Country Y Date 243,692 Switzerland Jan,16, 1947

